The present invention relates to a two-stroke cycle gasoline engine, and, more particularly, to a two-stroke cycle gasoline engine adapted for use with automobiles.
A two-stroke cycle engine has theoretically the advantage that an engine of a certains size can generate a greater power than a four-stroke cycle engine of a bigger size because the two-stroke cycle engine has twice as many work cycles per revolution as the four-stroke cycle engine. In fact, however, the conventional two-stroke cycle gasoline engine employing a carburetor has such drawbacks that it has high fuel consumption as compared with the four-stroke cycle engine due to the loss of air-fuel mixture caused by the direct escape, i.e. blow-out, of scavenging mixture to an exhaust manifold during scavenging, and that it cannot generate such a high power as expected from the fact that it has twice as many work strokes as the corresponding four-stroke cycle engine, due to the fact that the scavenging is still insufficient. Because of these problems, the practical use of two-stroke cycle gasoline engines is nowadays limited to the field of small engines which must be simple in structure and low in manufacturing cost.
Conventional two-stroke cycle gasoline engines of the abovementioned type, therefore, generally employ crankcase compression for scavenging. However, the scavenging by crankcase compression is not fully effective and can only provide a relatively low volumetric efficiency. This is the principal cause of the poor output power of conventional two-stroke cycle gasoline engines. In fact, a volumetric efficiency as high as 80% is available in four-stroke cycle engines, while on the other hand the volumetric efficiency of typical two-stroke cycles engines is still as low as 40-50%. The pump stroke volume of crankcase compression is equal to the stroke volume of the engine. However, since the crankcase has a relatively large clearance volume, the compression ratio of crankcase compression is relatively low, so that as a result the amount of air-fuel mixture drawn to the crankcase is small, the amount of delivered mixture is small, the delivery pressure is low and hence the scavenging pressure is low, and consequently it is hard to supply a really adequate amount of scavenging mixture into the power cylinder. As a result, the delivery ratio obtained in an engine wherein scavenging is effected only by the normal crankcase compression is only as high as 0.5-0.8. Since further the trapping efficiency is about 0.7, the volumetric efficiency becomes as low as 40-50% as mentioned above.
The purpose of scavenging is to push the residual exhaust gases remaining in the power cylinder out of it by fresh mixture, and therefore if the pressure of the residual exhaust gases and the distance between the scavenging port and the exhaust port are given, the time required for completing scavenging is determined by the pressure and the amount of scavenging mixture, provided that stratified scavenging is performed. Now, if the scavenging pressure is low, as when crankcase compression is used, a relatively long time is required for completing scavenging, particularly when the scavenging is performed by uniflow scavenging. Therefore, when the engine is rotating at high speed, it may well occur that the exhaust port is closed before the scavenging is completed so that a large amount of exhaust gas still remains in the power cylinder, and thereby only a very small amount of fresh mixture is charged into the power cylinder. Therefore, conventional two-stroke cycle engines of uniflow scavenging type have been unable to operate satisfactorily in the high speed range.
Furthermore, when scavenging depends only upon crankcase compression, since a power piston also operates as a pump piston, as a matter of course, the operational phase difference between a power cylinder-piston assembly and a pump cylinder-piston assembly is exactly 180.degree.. Therefore, the pump piston of a pump cylinder-piston assembly just reaches its top dead center (TDC) when the power piston of a power cylinder-piston assembly reaches its bottom dead center (BDC). In this connection, in the present description the top dead center (TDC) of a piston means the dead center of the piston at the end of the compression stroke of the piston, while the bottom dead center (BDC) of a piston means the dead center of the piston at the end of the suction or expansion stroke of the piston. In this case, however, although a half of the scavenging period still remaining to a power cylinder-piston assembly when its power piston has reached its BDC, a pump piston now begins to move toward its BDC, whereby the pressure in the crankcase lowers to generate a partial vacuum in the crankcase while the scavenging port still opens after the power piston goes far beyond its BDC, thereby causing the problem that the scavenging period is not all effectively utilized.
In view of the aforementioned problem with regard to scavenging by crankcase compression, and noting that that problem is the principal reason why the conventional two-stroke cycle engine using crankcase compression cannot be an actually effective engine, we have proposed, in U.S. Pat. No. 4,185,596, a two-stroke cycle gasoline engine which aims at as high operational performance as possible according to the principle of a two-stroke cycle gasoline engine by replacing scavenging dependent upon crankcase compression with scavenging dependent upon a separate scavenging pump, and by simultaneously constructing a power cylinder-piston assembly to incorporate uniflow scavenging and two horizontally opposed pistons and by constructing the separate scavenging pump to be of a reciprocating type having total stroke volume of 1.15-1.65 times as large as that of the power cylinder-piston assembly, and by further incorporating an operational phase relation between the power and the pump cylinder-piston assemblies so that the top dead center of the pump cylinder-piston assembly is, as viewed in the crank angle diagram, in a range between 15.degree. in advance of and 15.degree. behind the midpoint between the bottom dead center and the scavenging port closing phase point of the power cylinder-piston assembly to which the pump cylinder-piston assembly supplies scavenging mixture.
However, although it provides a particular advantage to abolish the crankcase compression completely, as is the case with the two-stroke cycle gasoline engine proposed in the aforementioned patent application, on the other hand, it requires the separate scavenging pump to have a relatively large stroke volume thereby causing the overall dimensions of the engine to increase when compared with a corresponding engine which utilizes crankcase compression.
Therefore, it is an object of the present invention to provide a two-stroke cycle gasoline engine which produces a substantially higher power output per unit volume of the engine displacement than a conventional two-stroke cycle gasoline engine by incorporating a supplementary scavenging pump in addition to effectively utilizing the crankcase compression, thereby substantially increasing both the amount and the pressure of scavenging mixture. In this case, the present invention contemplates to combine scavenging by the crankcase compression and scavenging by the supplementary scavenging pump in a particular manner related with the operational phase of the engine.
In more detail, in accordance with the present invention, the supplementary scavenging pump operates in its initial stage of operation as a booster which supplies its delivery mixture to a crankcase so as to increase both the amount and the pressure of scavenging mixture supplied by crankcase compression. Subsequently, at a predetermined phase point in a phase region located before BDC of a power piston, the supply of scavenging mixture from the supplementary scavenging pump to the crankcase is interrupted, and thereafter, during the remaining period of the scavenging period, the scavenging mixture from the supplementary scavenging pump is directly supplied to the power cylinder, thereby maintaining effective scavenging flow in the power cylinder over a substantial part of the scavenging period in which the scavenging ports are opened. For this purpose, the operational phase of the supplementary scavenging pump is somewhat delayed relative to that of the crankcase compression, that is, the piston of the supplementary scavenging pump reaches its TDC after the power piston has passed its BDC.
Known methods of scavenging in two-stroke cycle engines are conventionally cross scavenging, loop scavenging, and uniflow scavenging. In this connection, and in connection with the aforementioned high pressure scavenging contemplated in the present invention, if the scavenging pressure is increased in cross or in loop scavenging, the flow of scavenging mixture is liable to penetrate through exhaust gases existing in the power cylinder in a short-cutting manner, and also scavenging mixture and exhaust gases may be mixed with each other thereby not only causing poor scavenging, but also increasing the above explained blow-out loss of mixture, thus lowering the volumetric efficiency. On the other hand, it has been experimentally confirmed that, when uniflow scavenging is employed, it is possible to push uniformly the exhaust gases existing in the power cylinder out of it by the scavenging mixture at high pressure without causing any detrimental mixing between the scavenging mixture and the exhaust gases, and that, in this case, if the amount of scavenging mixture is increased so as to be necessary and sufficient, and if the duration of actual supply of scavenging mixture is long enough, scavenging at high scavenging efficiency is accomplished, and, as a result, the volumetric efficiency increases, resulting in corresponding increase of engine output power.
Therefore, it is another object of the present invention to provide a two-stroke cycle gasoline engine having a high volumetric efficiency in which high scavenging pressure and long-lasting supply of scavenging mixture effected by the aforementioned particular combination of crankcase compression and a supplementary scavenging pump are combinedly incorporated with uniflow scavenging.
By the aforementioned combination of high pressure and long-lasting supply of scavenging mixture with uniflow scavenging, the two-stroke cycle gasoline engine of the present invention is able to operate even in the relatively high speed operational range in which the conventional crankcase compression type two-stroke cycle gasoline engine is unable to operate with sufficient output power, and, therefore, it is possible to increase further the output power per unit volume of the space which the engine occupies. However, it is to be noted that the aforementioned relatively high speed rotational region is located in a lower speed region than the high rotational speed region of conventional automobile four-stroke cycle gasoline engines, and, therefore, it is still another object of the present invention to provide a two-stroke cycle gasoline engine which can generate sufficient output power in such a lower speed rotational region. Conventionally, a relatively small-sized four-stroke cycle gasoline engine for automobiles is designed so as to be operated at relatively high rotational speed so that relatively high power output is available from a relatively small-sized engine. In this connection, it is noted that, for example, in the case of an engine which has a two liter piston displacement and produces 92 PS of brake horsepower at 5000 rpm, a very large proportion of the power, such as 52 PS out of the indicated horsepower of 144 PS, is consumed by internal friction losses in the engine. The ratio of the internal friction loss to the output power of the engine is substantially reduced by lowering the rotational speed of the engine. The fact that a two-stroke cycle engine can generate higher power than a four-stroke cycle engine at lower rotational speed if its volumetric efficiency is increased is attributed to the feature that a two-stroke cycle engine has twice as many work cycles per revolution as a four-stroke cycle engine. The present invention contemplates effective utilization of this feature by increasing the volumetric efficiency of a power cylinder so that the rotational speed of the engine may be lowered to have a sufficient power output and so that the net output power per unit stroke volume of the engine may be increased. The maximum rotational speed of the engine contemplated in the present invention is 3800 rpm at the highest.
Furthermore, currently there exists a great demand for the development of cars which have low fuel consumption, in view of energy saving. Cars also must satisfy a high standard with regard to the prevention of air pollution. In order to improve fuel consumption, not only improvement of the fuel consumption of the engine itself, but also reduction of the air resistance of the vehicle are required. We noted, in connection with various running tests carried out to prepare for the qualification test for conforming to the standards for the prevention of air pollution (which are becoming more severe nowadays), that fuel consumption is different in summer and in winter due to the difference of atmospheric air density, and we more keenly recognized that the air resistance of the vehicle has an important effect on the fuel consumption of the vehicle even in low speed running.
In order to lower the air resistance of the vehicle it is important to reduce the height of the vehicle as much as possible and to form the vehicle in a streamlined external shape. Particularly, it is very effective to lower the engine hood. In order to reduce the height of the vehicle it is effective to eliminate the drive shaft for driving the rear wheels so that the shaft tunnel is eliminated and a flat floor is available over the entire floor area, thereby constructing a vehicle body having a low floor and a low roof. A method for accomplishing this is to employ the FF system, i.e. the front engine-front drive system. In order to lower the engine hood by a large amount in an automobile of FF type while ensuring necessary legroom for the driver and the front seat passenger, it is necessary to reduce substantially the height and length of the engine compartment. Furthermore, in order to reduce the air resistance of the vehicle, it goes without saying that the frontal area of the vehicle must be reduced. Therefore, the width of the vehicle should be minimized. Furthermore, since the transmission, differential gears, and other driving mechanisms must be housed in the engine compartment together with the engine, in the FF system, the space allowed for the engine is much reduced. Light trucks are often designed with the engine mounted under the driver's seat, and in such a design the engine, being relatively long, often extends so far backwards as to make a hump formed of the engine enclosure rearward of the cabin, thus shortening the deck.
It is, therefore, still another object of the present invention to provide a small-size gasoline engine having a low height, a small length and not a very large width, yet being capable of generating high power relative to the size of the engine. When this object is considered together with the other objects of the present invention which should be accomplished by a two-stroke cycle gasoline engine incorporating uniflow scavenging, it is contemplated that it is the most advantageous that the two-stroke cycle gasoline engine incorporating uniflow scavenging should employ two horizontally opposed pistons structure, as the power assembly, although other types of uniflow scavenging power assemblies, such as an assembly having a poppet valve in an exhaust port, are available.
Based upon the aforementioned basic structure, the present invention further proposes to determine the stroke volume of the supplementary scavenging pump to be 0.35-0.85 times as large as the total stroke volume of the power cylinder-piston assembly to which the pump cylinder-piston assembly supplies scavenging mixture. In this case, the total stroke volume of the power cylinder-piston assembly means the stroke volume displaced by the power piston when it moves from its BDC to its TDC. Therefore, the net compression stroke volume displaced by the power piston after it has closed the exhaust port until it reaches its TDC is smaller than the total stroke volume. When an engine includes more than one power cylinder or pump cylinder, the total stroke volume of the power cylinders or the pump should be the sum of the total stroke volume of applicable cylinders of the power cylinders or pump. In any event, in accordance with the present invention, it is proposed that the total stroke volume of the supplementary scavenging pump should be 0.35-0.85 times as large as the total stroke volume of the power cylinder to which the supplementary scavenging pumps supply scavenging mixture. The numerical limitation of 0.35-0.85 times as large has been determined in consideration of the amount and the pressure of scavenging mixture required for accomplishing high volumetric efficiency of supply of fresh mixture to a power cylinder such as 75%-100% or more in some cases.
In more detail, in order to increase volumetric efficiency of supply of fresh mixture to a power cylinder, the supply of scavenging mixture should be increased. However, excessive increase of the supply of scavenging mixture causes blow-out of scavenging mixture to the exhaust manifold during scavenging and increases fuel consumption and the density of hydrocarbons in the exhaust gases. The aforementioned numerical value of 0.85 is an upper limit determined in consideration of the fact that the engine of the present invention is a gasoline engine. On the other hand, the numerical value of 0.35 is a lower limit which may be used when the volumetric efficiency of the power cylinder needs to be only about 75%.
The time required for scavenging a power cylinder is determined by various factors such as the pressure difference between the pressure of scavenging mixture and that of the exhaust gases existing in the power cylinder when the scavenging port is opened, the length of the path along which scavenging mixture flows in the power cylinder while pushing exhaust gases until it reaches the exhaust port, i.e. the length of the helical path of scavenging mixture defined in the power cylinder in the case of the present invention which employs uniflow scavenging taking helical scavenging paths, and the behaviour of the initial and subsequent flow of scavenging mixture after the opening of the scavenging port, i.e. the initial pressure and the initial flow rate of scavenging mixture and subsequent changes of the pressure and the flow rate, and is not directly related with the rotational speed of the engine. However, the period which affects effectiveness of scavenging is the period between the opening and the closing of the scavenging port and since this period is shorter as the rotational speed of the engine is higher, it is necessary that the pressure and the amount of scavenging mixture should correspondingly increase as the rotational speed of the engine increases so that scavenging is completed before the scavenging port is closed. In this case, it is required that the introduction of scavenging mixture to the power cylinder should be maintained for a substantial time duration from the moment of opening of the scavenging port. Scavenging pressure available in the conventional crankcase compression is 0.3-0.35 atm (gauge pressure) at the highest due to a relatively large clearance volume involved therein. In the case of a particular modern design of crankcase it is possible to obtain scavenging pressure of about 0.45 atm. Even in this case, however, the scavenging pressure substantially lowers before the power piston reaches its BDC, and the amount of scavenging mixture is still insufficient.
In accordance with the present invention, when the engine is designed so as to accomplish sufficient scavenging in high speed operation of the engine, the scavenging pressure is increased to be above 0.4 atm and such a high scavenging pressure is maintained for a substantial time in the scavenging period.
A supplementary scavenging pump having total stroke volume 0.35-0.85 times as large as that of the power cylinder-piston assembly to which it supplies scavenging mixture should, in connection with the feature that the power cylinder-piston assembly is constructed so as to have horizontally opposed pistons, preferably be a pump cylinder-piston assembly having two horizontally opposed pistons. In this case, it is possible to harmonize the diameter of the pump cylinder and the stroke of the pump piston with those of the power cylinder-piston assembly, and it also provides the convenience with regard to the driving connection of the power cylinder-piston assembly and the pump cylinder-piston assembly that a pair of crankshafts arranged at opposite ends of the power cylinder-piston assembly need only to be extended so as also to serve as a pair of crankshafts of the pump cylinder-piston assembly.
In this connection, if a pair of crankshafts in the aforementioned driving structure are drivingly connected with each other with interposition of a rotational direction reversing means such as an idler gear, balancing of the rotational moment of the crankshafts and the power and pump pistons will be obtained. However, a rotational direction reversing means which includes an idler gear, etc. needs additional devices and will increase the manufacturing cost of the engine. Therefore, from cost-benefit evaluation the present invention contemplates, as an embodiment thereof, to connect a pair of crankshafts for the associated power and pump cylinder-piston assemblies each having two horizontally opposed pistons simply by an endless chain so that the two crankshafts are rotated in the same direction. However, it is a matter of balance between pursuit of silence and pursuit of low cost to select between co-rotation or counter-rotation of a pair of crankshafts, and it is a matter of design to be determined in view of design conditions of the engine.
Although it has been mentioned that the supplementary scavenging pump to be incorporated in a two-stroke cycle gasoline engine of the present invention should preferably be a pump cylinder-piston assembly having two horizontally opposed pistons, of course, the supplementary scavenging pump may be a pump cylinder-piston assembly having a single pump piston. When a single piston type pump cylinder-piston assembly is employed, there is the disadvantage on the one hand that relative dimensions and harmony between the power cylinder-piston assembly and the pump cylinder-piston assembly and dynamic balance of the pump cylinder-piston assembly are less desirable than when a pump cylinder-piston assembly having two horizontally opposed pistons is used, while on the other hand there is the advantage that the manufacturing cost is reduced. Particularly for small sized engines, the advantage of reduction of manufacturing cost may well get the better of the abovementioned structural disadvantage.
As mentioned above, one of the basic constitutions of the present invention is the combination of a supplementary scavenging pump with the conventional crankcase compression so as to maintain high pressure and flow of scavenging mixture for a relatively long time after the scavenging port has been opened, thereby accomplishing high volumetric efficiency of the power cylinder. For this purpose, in accordance with the present invention, it is proposed that the top dead center of the supplementary scavenging pump of the reciprocating type should be delayed relative to the bottom dead center of the power cylinder-piston assembly, or, in other words, to delay the bottom dead center of the supplementary scavenging pump relative to the top dead center of the power cylinder-piston assembly, i.e., the bottom dead center of the crankcase compression.
However, if this delay in phase is excessive, the delivery pressure of the supplementary scavenging pump at the time point when the scavenging port is opened will become insufficient, and, therefore, the pressure in the crankcase will not be increased up to the level which is required for accomplishing the effect intended by the present invention. Particularly when it is intended to increase volumetric efficiency of an engine in its high speed operation, high scavenging pressure is demanded at the time point when the scavenging port is opened. In this case, therefore, it is required that the supplementary scavenging pump should have covered a substantial phase angle from its bottom dead center before the scavenging port is opened. In consideration of this factor, the BDC of the supplementary scavenging pump is determined, as shown in FIG. 1, to be, as viewed in the crank angle diagram, in a range between 15.degree. in advance of and 15.degree. behind the point which is 90.degree. in advance of the scavenging port opening phase point (So) of the power cylinder-piston assembly to which the supplementary scavenging pump supplies scavenging mixture. By positioning BDC of the supplementary scavenging pump in such a phase range and by conducting the delivery mixture of the pump into the crankcase, the pressure in the crankcase is properly increased so that the scavenging pressure in the initial stage of scavenging period is desirably increased by co-operation of the crankcase compression and the supplementary scavenging pump to such an extent as effectively to accomplish the purposes of the present invention. In this case, the scavenging pressure at the scavenging port opening phase point (So) is determined by various factors such as: clearance volume of the crankcase; stroke volume of the supplementary scavenging pump; crank angle between BDC of the supplementary scavenging pump and the scavenging port opening phase point So; whether the delivery mixture of the supplementary scavenging pump is supplied to both crankcases of a power cylinder-piston assembly having two horizontally opposed pistons or to only one of the two crankcases; when the scavenging mixture is supplied to only one crankcase, whether or not the scavenging mixture from said one crankcase interferes with the scavenging mixture from the other crankcase so as to reduce its pressure; etc..
However, even when the pressure and the flow of scavenging mixture delivered from the crankcase are increased by introducing scavenging mixture delivered from the supplementary scavenging pump to the crankcase, when the power piston approaches its BDC, the pressure in the crankcase abruptly lowers because the movement of the power piston is substantially slowed down as it approaches its BDC, thereby reducing the effect of crankcase compression, and furthermore the crankcase enters into its suction stroke after the power piston has passed its BDC. In view of this, the present invention contemplates to interrupt the supply of scavenging mixture from the supplementary scavenging pump to the crankcase before the power piston reaches its BDC, and to supply thereafter the scavenging mixture delivered from the supplementary scavenging pump directly to the power cylinder. By this arrangement, the supply of scavenging mixture to the power cylinder is effectively maintained even in the latter half of the scavenging period, i.e. after BDC of the power piston, and it is avoided that the scavenging mixture delivered from the supplementary scavenging pump should be uselessly supplied to the crankcase which is performing its suction stroke. In this case, the closer the BDC of the supplementary scavenging pump is approached to the scavenging port opening phase point So, or the earlier the supply of scavenging mixture from the supplementary scavenging pump to the crankcase is interrupted, the more scavenging mixture remains in the supplementary scavenging pump at the time of interruption, so that a larger amount of scavenging mixture is directly supplied from the supplementary scavenging pump to the power cylinder after the interruption of the supply of scavenging mixture from the supplementary scavenging pump to the crankcase. In accordance with the present invention, the time point of the interruption, i.e. the time point when the passage which conducts scavenging mixture from the supplementary scavenging pump to the crankcase is interrupted, is determined, as viewed in the crank angle diagram, to be in a range 10.degree.-30.degree. before BDC of the power piston. In this connection, in the present specification the opening phase point with regard to the scavenging or exhaust port respectively means the phase point at which the scavenging or exhaust port begins to open, and the closing phase point with regard to the scavenging or exhaust port respectively means the time point at which the scavenging or exhaust port begins to close, in the unsual manner of using these terms.
In connection with the aforementioned feature of the present invention that the supply of scavenging mixture from the supplementary scavenging pump to the crankcase is interrupted part way through the scavenging period and thereafter the scavenging mixture delivered from the supplementary scavenging pump is directly supplied to the power cylinder, the present invention contemplates, as another additional feature, to provide the passage for directly supplying scavenging mixture from the supplementary scavenging pump to the power cylinder to be substantially throttled at its discharge end towards the inside of the power cylinder. Nowadays it is practiced to increase the air/fuel ratio of the mixture supplied to a gasoline engine, i.e. to employ lean mixture, in order to reduce the amount of harmful components in exhaust gases. Such lean combustion is particularly employed in partial load operation of the engine. In partial load operation the throttle valve is partly closed, and as a matter of course the delivery ratio of scavenging mixture is low, thereby causing an increased amount of exhaust gases to remain in the power cylinder. Therefore, in partial load operation incorporating lean combustion, the ignitability of the fuel-air mixture lowers greatly, and the burning velocity of the fuel-air mixture also greatly lowers. In accordance with the present invention, these problems are effectively obviated by the scavenging mixture directly supplied from the supplementary scavenging pump to the power cylinder from part way through the scavenging period being ejected as a jet into the power cylinder due to the aforementioned throttling of the passage which conducts scavenging mixture directly from the supplementary scavenging pump to the power cylinder in its discharge end opened toward the inside of the power cylinder, whereby the jet generates turbulences in the power cylinder which effectively improve the ignitability and burning velocity of the fuel-air mixture. Since in this case such turbulences of scavenging mixture are generated principally in the latter half portion of the scavenging period, disadvantages such as causing of mixing between scavenging mixture and exhaust gases in the power cylinder or such as causing direct blow-out of scavenging mixture to the exhaust manifold during scavenging, which would occur if such turbulences were generated in an early stage of scavenging, are not suffered. Furthermore, since the turbulences of scavenging mixture due to the aforementioned jet injection of mixture are generated after BDC of the power piston, the turbulences are maintained during the compression stroke of the power cylinder.
When a gasoline engine of the present invention includes, for example, two power cylinder-piston assemblies each having two horizontally opposed pistons, the supplementary scavenging pump may be of either of the following two types. In one type, with the two power cylinder-piston assemblies being operated in exactly the same phase, the supplementary scavenging pump is a single acting piston pump having total stroke volume twice as large as that required for one power cylinder-piston assembly. In this case, if the stroke volume which the supplementary scavenging pump should have for one power cylinder-piston assembly is relatively small so that a one-piston single acting pump is able to have the required stroke volume, it is desirable that the supplementary scavenging pump should be a pump cylinder-piston assembly having two horizontally opposed pistons so that it provides a double stroke volume required to serve for two power cylinder-piston assemblies. On the other hand, if the power cylinder-piston assembly is relatively large so that one pump cylinder-piston assembly having two horizontally opposed pistons is required for one power cylinder-piston assembly, it is desirable that the pump cylinder-piston assembly should be modified without changing the stroke of the pump pistons to have an increased diameter of pump cylinder to be approximately the square root of two times as large. In the other type, with the two power cylinder-piston assemblies being operated with a phase difference of 180.degree. from each other, the supplementary scavenging pump is of a double acting type. In this case, torque fluctuation of power output shaft becomes small.